1. Field of the Invention
The present invention relates to a drum motor for a VCR (Video Cassette Recorder; a cassette VTR) which employs hydrodynamic bearings as radial and thrust bearings. More particularly, the present invention relates to a drum motor for a VCR which rotates with minimal vibration, satisfies the demand for higher recording densities of a VCR and has excellent durability and reliability.
2. Prior Art
VCR systems for business use are now shifting from analog recording to digital recording, to attain higher recording densities. As one of the measures taken to increase recording density, it is common practice to narrow the width and pitch of tracks recorded on a magnetic tape. To realize this, a technique of minimizing the run-out of the rotor-side unit is indispensable.
FIG. 9 is a fragmentary sectional view of a conventional drum motor for a VCR. Referring to FIG. 9, a rotating shaft 104 is rotatably supported by the inner peripheral surface 101a of the bottom center of a motor casing 101 and the inner peripheral surface 101b of the top center or the motor casing 101 through ball bearings 102 and 103. A flange 105 is secured to the outer periphery of the upper end of the rotating shaft 104 so that a rotor-side unit 110 can be mounted thereon. A rotor magnet 106 of a motor part for driving the rotating shaft 104 to rotate is coaxially secured to the outer periphery of the central portion of the rotating shaft 104 through a cylindrical back yoke 107 made of a magnetic material or the like. A stator core 108 is secured to the inner peripheral surface of the motor casing 101 in an opposing relation to the rotor magnet 106.
Incidentally, in the above-described drum motor, a stator-side unit 109 is generally secured to the top of the motor casing 101, whereas the rotor-side unit 110 is secured to the flange 105 to constitute a drum unit. A magnetic tape for recording is wound on both the outer peripheral surface 109a of the stator-side unit 109 and the outer peripheral surface 110a of the rotor-side unit 110, and in this state, the tape is fed at a constant speed so that recording tracks are provided on the magnetic tape by means of a magnetic head 111 attached to the rotor-side unit 110a.
In the conventional drum motor employing ball bearings, however, the magnitude of vibration of the rotating shaft 104 depends on the bearing clearance of the ball bearings 102 and 103. The vibration in the radial direction is substantially equal to the radial clearance of the ball bearings 102 and 103, while the vibration in the thrust direction is substantially equal to the thrust clearance of the ball bearings 102 and 103. To minimize the bearing clearance, various schemes have been devised, for example, application of a preload to the ball bearings 102 and 103. However, the non-repeatable run out, i.e. the non-repeatable component of vibration in the radial and axial directions, which cannot be removed even after a coaxial machining of the flange to remove vibration thereof, is of the order of 0.3 to 0.5 microns, and the minimization of the vibration of the rotating shaft 104 has already reached the limit due to problems peculiar to ball bearings. Thus, the vibration of the rotating shaft 104 is a serious obstacle to the achievement of a higher recording density on magnetic tape.
On the other hand, as a device for realizing a rotational performance of high accuracy, there is a drum motor that employs hydrodynamic bearings, as shown in Japanese Patent Application Post-Exam Publication No. 63-64645. FIG. 10 schematically shows the structure of this drum motor. Referring to FIG. 10, a disc 123 is rotatably fitted on a fixed shaft 121, and a thrust bearing member 124 is attached to the top of the disc 123. The fixed shaft 121 is provided at two portions thereof with radial grooves 126A and 126B, which constitute a radial hydrodynamic bearing in combination with the inner peripheral surface of the disc 123. In addition, the end surface of the fixed shaft 121 and the thrust bearing member 124, which faces it, constitute in combination a thrust hydrodynamic bearing. Thus, the disc 123 is rotatably supported on the fixed shaft 121 without coming into solid contact with it. Accordingly, it is possible to rotate the disc 123 smoothly with minimal vibration and high accuracy in comparison to the conventional drum motor employing ball bearings.
However, it is necessary in order to narrow the recording track pitch, as described above, to reduce the run-out of the rotor-side unit. In particular, vibration in the thrust direction and a change in the relative height between the rotary portion and the stationary portion in the thrust direction directly obstruct the achievement of a reduction in the track pitch.
Further, with the arrangement of the thrust hydrodynamic bearing in the prior art shown in FIG. 10, the thrust load bearing direction is limited to one. Therefore, the force that is produced in the thrust hydrodynamic bearing during the rotation of the rotating shaft is equal to the thrust load and will never exceed it. Further, the amount of levitation of the thrust bearing is a function of the hydrodynamic force and the thrust load. Therefore, if the thrust load changes due to a change of the position or attitude in which the motor is used, the amount of thrust levitation changes substantially according to the change in the thrust load. For example, when the motor is used in a horizontal position, the thrust load is approximately zero, and the amount of thrust levitation is infinite. In an extreme case, the rotor-side unit may fall off and break down.
To overcome these disadvantages, the conventional drum motor is arranged such that a change in the amount of levitation of the thrust bearing due to a change of the position in which the motor is used, that is, a change in the relative height of the rotor-side unit relative to the stationary-side unit in the thrust direction, is reduced by using the axial attraction force of a permanent magnet 138, thereby enabling the motor to be used even in a horizontal position. However, in the conventional drum motor shown in FIG. 10, the permanent magnet 138 is disposed to face an iron piece 142 to exert an attraction force in the downward direction, i.e., the direction of the arrow X. Accordingly, when the motor is used in a vertically upward position, the attraction force of the permanent magnet 138 and the weight of the rotor act in the same direction, and the thrust load is the sum of the attraction force of the permanent magnet and the weight of the rotor, which increases the wear of sliding surfaces of the bearing members.
Further, in general, in order to minimize the amount of levitation of the rotor when the motor is used in a horizontal position, it is necessary to reduce the difference between the thrust load of the motor when used in a horizontal position and the thrust load of the motor when used in a vertically upward position. However, in the conventional drum motor shown in FIG. 10, when the motor is used in a horizontal position, the thrust load is only the attraction force of the permanent magnet, and the rate .epsilon. of relative change of the thrust load is expressed by ##EQU1##
Therefore, in order to minimize the relative change of the thrust load to make the change rate .epsilon. in the above expression approach 1, it is necessary to increase the attraction force of the permanent magnet relative to the weight of the rotor.
It is reported that in the prior art shown in FIG. 10 the change in the relative height of the rotor-side unit due to a change of the position in which the motor is used has been reduced to 0.5 microns by setting the attraction force of the permanent magnet at a level double the weight of the rotor (i.e., .epsilon.=0.67). With the bearing structure shown in this prior art, however, normal operation cannot be expected when the motor is used in an inverted position. Namely, in an inverted position of the motor, the numerator of the above expression is expressed as attraction force of permanent magnet - (minus) weight of rotor. Hence, it is necessary in order to make the change rate .epsilon. approach 1 to further increase the attraction force of the permanent magnet. Consequently, an excess thrust load is applied to the thrust bearing.
An excess thrust load causes an increase in the starting torque of the motor and also accelerates wear on the bearing sliding surfaces due to the repetition of start and stop, giving rise to problems in terms of durability. Further, it may be inferred that the prior art shown in FIG. 10 gives no consideration to the normal operation of the motor in an inverted position from the fact that it needs a stopper ring 139 for preventing the rotor-side unit from falling off when the motor is in an inverted position.
Further, in the above-described prior art, the thrust bearing is composed of the end surface of the fixed shaft 121 and the thrust bearing member 124, which faces it. Therefore, the outer diameter of the thrust dynamic pressure generating surface cannot be increased as desired, and hence the thrust load bearing capacity is low relative to the radial load bearing capacity. Accordingly, a high-viscosity fluid, e.g., lubricating oil, is used in this type of hydrodynamic bearing because if a non-viscosity fluid like air is employed, the load bearing capacity, particularly that of the thrust bearing is insufficient in general. However, the viscosity of lubricating oil shows a logarithmic change with temperature and becomes particularly high at low temperatures. This implies that a temperature change causes a change in the amount of levitation of the thrust bearing, leading to a change in the relative height of the rotor-side unit and other problems related thereto.
There are also problems in terms of reliability, for example, degradation of the performance caused by a drying up of lubricating oil due to scattering or leakage thereof or contamination of the parts surrounding the bearing by lubricating oil, and other problems, for example, an increase in the motor current due to an increase in the bearing torque at low temperatures.
Under these circumstances, an H-shaped bearing structure similar to the bearing of the drum motor according to the present invention, which has two thrust plates, has been proposed as disclosed in the specification of U.S. Pat. No. 3,950,039. FIG. 11 schematically shows the H-shaped bearing structure. Referring to FIG. 11, two thrust plates 203 are brought into direct contact with both end surfaces of a radial bearing member 205, and these plates and the bearing member are secured to a fixed shaft 204 by tightening a nut 208, for example. In the H-shaped bearing structure, the two thrust plates 203 are provided with respective dynamic pressure generating parts 209 such that the directions of dynamic pressures generated are opposite to each other.
Thus, the bearing rigidity in the thrust direction is determined by clearances at both ends of the thrust bearing. By reducing the clearances, the bearing rigidity can be enhanced, and it is possible to realize a hydrodynamic bearing having high rigidity even if air is employed as a fluid. In addition, it is possible to minimize the change of the position of the rotor in the thrust direction due to a change of the position of the motor in which it is used, and it is possible to eliminate the need for the attraction force in the thrust load direction by the permanent magnet as shown in FIG. 10.
However, a bearing having the structure shown in FIG. 11 also suffers from problems. For example, if an improper torque is applied for tightening the nut 208, the thrust plates 203 are deformed. As a result, the expected uniform clearances cannot be ensured, and the expected bearing performance cannot be obtained. To cope with such problems, attempts have been made to minimize the deformation of the thrust plates 203, for example, by setting the tightening torque at a level which is a fraction of the standard value (i.e., the torque at which the tensile force arising from the axial force caused by the bolt tightening torque reaches 70% of the yield point of the material concerned). With this method, however, it is impossible to sufficiently ensure the force required to secure the rotor-side bearing member to the rotating shaft. Accordingly, when the shaft-side members are used as a rotor in the above-described H-shaped bearing structure, reliability during high-speed rotation is unsatisfactory in view of the influence of centrifugal force acting on the bearing during rotation.
In a case where a rotating shaft 5 is provided with a flange 10 for mounting a rotor-side unit, e.g., a recording head, a rotary drum, etc., as shown in FIG. 12, since the rotor-side unit is mounted directly on the flange 10, the accuracy of rotation of the flange 10, i.e. coaxial rotation of the flange and the rotary shaft, directly determines the accuracy of rotation of the rotor-side unit. Therefore, it is necessary in order to realize higher recording densities of VCR to raise the rotational accuracy of the flange 10. To minimize the run-out of the flange 10 during rotation, it is a general practice to machine the rotor-side unit mounting surface 10b of the flange 10 after it has been secured to the rotating shaft 5 so that the unit mounting surface 10b is perpendicular to the rotational axis of the rotating shaft 5 (coaxial machining), thereby obtaining a desired degree of rotational accuracy.
However, the flange 10 is secured to the rotating shaft 5 by means, for example, of a locking member, e.g., a locking nut 11 which clamps the center portion of the flange. In this case, the machined surface 10b is deformed with time, as shown by a2 in FIG. 12, even after the flange 10 has been machined coaxially, by the clamping stress, i.e. uneven residual stress in the flange, imposed by the locking nut 11, for example, thus giving rise to problems such as an increase in the run-out of the machined surface 10b. In addition, the clamping stress exerted by the locking nut 11 causes deformation of a stepped portion 5a provided on the rotating shaft 5 and further causes a rotor-side bearing member 17 to be deformed as shown by a1 in the figure. To prevent such deformation, it has been a conventional practice to lengthen the axial dimension l1 of the stepped portion 5a to thereby increase the rigidity. However, an increase in the axial dimension l1 results in an increase in the overall length of the rotating shaft 5. Thus, such a conventional practice is an obstacle to a reduction in the overall size of the motor.
Although attempts have been made to minimize the deformation of the flange 10 and the rotor-side bearing member 17 by setting the tightening torque of the locking nut 11 at a level which is one over several (in a fractional expression) of the standard value, the force required to secure the flange 10 cannot sufficiently be ensured by this method, and satisfactory reliability cannot be obtained. In many cases, the flange 10 is made of the same material as that used to form a rotary drum or other member mounted thereon, e.g., an aluminum alloy, because the coefficient of linear expansion of the flange 10 needs to be coincident with that of the member mounted thereon.
On the other hand, the rotating shaft is formed by using stainless steel or other similar material in many cases, although it is different in the coefficient of linear expansion From a material used for the bearing members, e.g., a ceramic material, by taking into consideration the rigidity and workability of such material. Thus, the rotor is composed of a combination of different kinds of materials, which have different coefficients of linear expansion, and it is likely that the bearing members and the flange will be deformed or displaced at the joint due to a change in temperature. Accordingly, it is extremely difficult to maintain rotation of high accuracy over the entire range of temperatures at which the motor is used.